Heat exchanger and vehicle air conditioner equipped with the same

ABSTRACT

Provided is a heat exchanger whose air pressure loss is small and whose exchanged heat is large. The heat exchanger includes flat tubes, each having therein a plurality of refrigerant circulation holes, and fins fixed to the flat surfaces of the flat tubes, wherein the flat tubes and the fins are alternately stacked to form the heat exchanger.

TECHNICAL FIELD

The present invention relates to a heat exchanger and a vehicle air conditioner equipped with the same.

BACKGROUND ART

A vehicle air conditioner is provided with a condenser (heat exchanger) that condenses refrigerant by exchanging heat with air. A condenser of a typical multiflow type that is frequently used is configured such that flat tubes and corrugated fins are alternately stacked. The flat tubes have a plurality of refrigerant circulation holes therein, and the protrusions of the corrugated fins are fixed to the flat surfaces of the flat tubes to allow air to pass over the surfaces of the corrugated fins.

Conceivable ways to improve the performance of this type of condenser include reducing the fin pitch and reducing the size of refrigerant circulation holes formed in the flat tubes.

However, reducing the fin pitch will increase the pressure loss of air passing therethrough, thus causing an increase in motor input power of a condenser-radiator fan module (CRFM). Reducing the size of the refrigerant circulation holes will increase the refrigerant pressure loss, thus causing an increase in the motive power of a compressor that compresses the refrigerant.

PTL 1 below discloses a multiflow-type refrigerant condenser designed to obtain maximum heat radiation performance in consideration of both airflow resistance and tube pressure loss. Specifically, PTL 1 defines the configuration of the heat exchanger with the relationship between an air vent opening ratio Pr (=Th/Tp) and the outer circumferential tube thickness Td by using the parameters Th, which is the height of each of the tubes in the stacking direction, Tr, which is the height of each of refrigerant passages in the tubes in the stacking direction, Td, which is the outer circumferential tube thickness between the outer surface of the tube and the refrigerant passage, and Tp, which is the pitch of the flat tubes in the stacking direction.

CITATION LIST Patent Literature

{PTL 1} The Publication of Japanese Patent No. 3922288 (Claim etc.)

SUMMARY OF INVENTION Technical Problem

However, PTL 1 does not consider the size of the refrigerant condenser in the airflow direction (in the widthwise direction of the tubes) at all That is, it does not rigorously examine changes in state quantity (for example, airflow resistance and heat exchange in the airflow direction) due to air passing through the fins. Accordingly, the performance of a heat exchanger that depends on the airflow resistance cannot be rigorously evaluated using the specifications of the refrigerant condenser configuration described in the literature.

The present invention is made in consideration of such circumstances, and an object thereof is to provide a heat exchanger whose air pressure loss (airflow resistance) is small and whose exchanged heat is large, as well as a vehicle air conditioner equipped with the same.

Solution to Problem

To solve the above problem, the heat exchanger of the present invention and the vehicle air conditioner equipped with the same adopt the following solutions.

A heat exchanger according to a first aspect of the present invention comprises flat tubes, each having therein a plurality of refrigerant circulation holes, and fins fixed to the flat surfaces of the flat tubes, on the surfaces of which air passes, wherein the flat tubes and the fins are alternately stacked to form the heat exchanger, wherein (W−t1−t2)×Hp×Hf/N is set to 3.95 or greater and 10.0 or less, where de is the equivalent diameter of the plurality of the refrigerant circulation holes, W is the width of each of the flat tubes, t1 is the wall thickness of each of the flat tubes at one end in the widthwise direction corresponding to the distance from the refrigerant circulation hole closest to the one end, t2 is the wall thickness of each of the flat tubes at the other end in the widthwise direction corresponding to the distance from the refrigerant circulation le closest to the other end, Hp is the height of each of the flat tubes in the stacking direction, Hf is the height of the fins in the stacking direction, and N is the number of the refrigerant circulation holes, and the equivalent diameter de is set to 0.5 or more and 0.8 or less, and the flat tube width W is set to 12 mm or more and 16 mm or less.

As expressed in the polynomial, the polynomial using the flat tube width W was used to evaluate the heat exchange performance of the heat exchanger. This allows also state changes due to air passing through the fins (for example, airflow resistance and heat exchange in the airflow direction) to be considered, thus allowing the heat exchange performance to be reflected more rigorously.

The configuration of the heat exchanger is determined in consideration of not only the configuration of the flat tubes, such as the flat tube width W and the flat tube height Hp, but also the fin height Hf. This allows the air pressure loss to be considered more rigorously.

Dividing the polynomial by the number of refrigerant circulation holes, N, allows evaluation of the performance per refrigerant circulation hole.

It was found that, using the above polynomial and setting the value to 3.95 or greater and 10.0 or less allows a heat exchanger whose air pressure loss is small and whose exchanged heat is large to be realized.

Preferably, the flat tubes are manufactured by extrusion.

Preferably, the equivalent diameter de is set to 0.55 or more and 0.76 or less.

Setting the equivalent diameter de to 0.55 or more and 0.76 or less can further decrease the air pressure loss and increase the exchanged heat.

Preferably, the fins have a corrugated shape whose pitch is set to 1.6 mm or more and 2.0 mm or less.

Setting the fin pitch to 1.6 mm or more and 2.0 mm or less can further decrease the air pressure loss and increase the exchanged heat.

A vehicle air conditioner according to a second aspect of the present invention is equipped with any one of the above heat exchangers.

By providing the foregoing heat exchanger, a high-performance vehicle air conditioner can be provided. The heat exchanger of the present invention is suitable for use as a condenser of a vehicle air conditioner.

Advantageous Effects of Invention

According to the present invention, a heat exchanger whose air pressure loss is small and whose exchanged heat is large, as well as a vehicle air conditioner equipped with the same, can be provided.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a front view of a condenser of a vehicle air conditioner according to an embodiment of the present invention.

FIG. 2 is a transverse cross-sectional view of a flat tube in FIG. 1.

FIG. 3 is a front view of corrugated fins in FIG. 1.

FIG. 4 is a graph showing the simulation results of the condenser shown in FIG. 1 plotted against a polynomial.

FIG. 5 is a graph showing the simulation results of the condenser shown in FIG. 1 plotted against an equivalent diameter.

DESCRIPTION OF EMBODIMENT

An embodiment according to the present invention will be described hereinbelow with reference to the figures.

FIG. 1 shows a front view of a condenser (heat exchanger) 1 according to this embodiment. The condenser 1 cools a high-temperature, high-pressure superheated gas refrigerant discharged from a compressor (not shown) in a refrigeration cycle of a vehicle air conditioner to condense it. The condenser 1 is disposed in the frontmost part in a vehicle engine compartment as a component of a condenser-radiator fan module (CRFM). An engine cooling radiator (not shown) and a cooling fan (not shown) are disposed in sequence behind the condenser 1. The condenser 1 is cooled by cooling air (outside air) blown by the cooling fan.

The condenser 1 includes a first header tank 11 and second header tank 12 pair disposed with a certain distance therebetween. The header tanks 11 and 12 are cylindrical in shape and are disposed with the longitudinal direction thereof oriented in the substantially vertical direction. A core portion 13 that performs heat exchange between air and the refrigerant is disposed between the header tanks 11 and 12.

The condenser 1 is of a multiflow type in which refrigerant flows through a plurality of parallel channels provided between the header tanks 11 and 12. The core portion 13 is equipped with flat tubes 14 extending horizontally between the header tanks 11 and 12 and corrugated fins 15 fixed to the flat surfaces of the flat tubes 14. The flat tubes 14 and the corrugated fins 15 are alternately stacked in the vertical direction to form the core portion 13.

FIG. 2 shows a transverse cross-section of the flat tubes 14. As shown in the figure, each flat tube 14 has therein a plurality of independent refrigerant circulation holes 20 formed in the longitudinal direction. The flat tubes 14 having the plurality of refrigerant circulation holes 20 can be manufactured by extruding a material made of an aluminum alloy.

The flat tubes 14 are connected to the first header tank 11 at one end in the longitudinal direction and are connected to the second header tank 12 at the other end. Thus, the refrigerant circulates between the header tanks 11 and 12 through the plurality of refrigerant circulation holes 20.

FIG. 3 shows a front view of the corrugated fins 15. As shown in the figure, the corrugated fins 15 have a corrugated shape. The corrugated fins 15 can be manufactured by pressing a plate material made of an aluminum alloy. The peaks 15 a and troughs 15 b of the corrugated fins 15 are joined to the flat surfaces of the flat tubes 14 by brazing. Air flows over the surfaces of the corrugated fins 15 to accelerate heat exchange between the air and the refrigerant.

As shown in FIG. 3, the height of the corrugated fins 15 is Hf, and the fin pitch is Pf.

As shown in FIG. 1, the interior of the first header tank 11 is separated into two chambers 17 and 18 by a separator 16, and the gas refrigerant coming from the compressor is introduced into the upper first chamber 17. The gas refrigerant flows into the second header tank 12 via the upper flat tubes 14 communicating with the first chamber 17, makes a U-turn in the second header tank 12, and thereafter flows into the lower second chamber 18 via the remaining lower flat tubes 14. The gas refrigerant is cooled and condensed by exchanging heat with air passing through the space between the flat tubes 14, so that the refrigerant becomes a gas-liquid two-phase flow in the refrigerant circulation holes 20 of the flat tubes 14 as the refrigerant condenses.

Next, test results of the heat exchange performance of the condenser 1 with the above configuration in a numerical simulation will be described.

As an index of the heat exchange performance, this embodiment adopts Q/Fa/APa corresponding to exchanged heat Q [W] relative to frontal surface area Fa [m²] and air pressure loss ΔPa [Pa]. Adopting this heat exchange performance index allows the pressure loss of air passing through the condenser 1 (specifically, the corrugated fins 15) to be taken into account. That is, the larger the exchanged heat Q and the smaller the air pressure loss ΔPa are, the larger the value the index takes.

The exchanged heat Q and the air pressure loss ΔPa are calculated using the following relational expression:

ΔPa=A×Pf ^(B)

Q=C×exp(−D×Pf)

where Pf is the fin pitch (see FIG. 3), and A, B, C, and D are constants.

The simulation also considers the pressure loss of the refrigerant flowing in the refrigerant passage holes 20 of the flat tubes 14. Specifically, the refrigerant pressure loss is calculated from the tube friction coefficient of the refrigerant circulation holes 20, the physical properties of the gas refrigerant and the liquid refrigerant, and so on. If the refrigerant pressure loss is large, the change in state quantity in a p (pressure)−h (enthalpy) chart of the refrigerant during heat exchange (while the refrigerant is condensed) shifts downward to the left from an ideal horizontal line (that is, the pressure and the temperature are constant), and the average temperature CTm of the refrigerant during condensation decreases. The decrease in average temperature CTm will decrease the exchanged heat Q, which is proportional to the difference between the average temperature CTm and the air temperature Tai (CTm−Tai). Accordingly, the smaller the refrigerant pressure loss is, the larger the exchanged heat Q becomes, and thus, the larger the value the foregoing heat exchange performance index will take.

The simulation adopts the following conditions:

inlet air temperature Tai=35° C.

inlet refrigerant pressure Pri=1.744 MPa

frontal air velocity Fvi=4.5 m/s

the degree of superheating at refrigerant inlet SH=20 K

the degree of supercooling at refrigerant outlet SC=10

fin pitch Pf=1.6 mm or more and 2.0 mm or less.

The following polynomial was used as a parameter that defines the configuration of the condenser 1:

(W−t1−t2)×Hp×Hf/N.

The variables in the above polynomial are as follows, as shown in FIG. 2:

W: the width of each of the flat tubes 14

t1: the wall thickness of each of the flat tubes 14 at one end in the widthwise direction corresponding to the distance from the refrigerant circulation hole 20 closest to the one end (the left end in FIG. 2)

t2: the wall thickness of each of the flat tubes 14 at the other end in the widthwise direction corresponding to the distance from the refrigerant circulation hole 20 closest to the other end (the right end in FIG. 3)

Hp: the height of each of the flat tubes 14 in the stacking direction (in the vertical direction)

Hf: the height of the corrugated fins 15 in the stacking direction (in the vertical direction)

N: the number of refrigerant circulation holes 20.

In the above polynomial, the reason why the wall thickness at one end, t1, and the wall thickness at the other end, t2, are subtracted from the flat tube width N is that heat exchange is substantially not performed in the region of the thicknesses t2 and t2.

The reason why the height of each of the flat tubes 14 in the stacking direction, Hp, and the height of the corrugated fins 15 in the stacking direction, Hf, are multiplied by the flat tube width W is that these parameters are proportional to the exchanged heat.

The expression is divided by the number of refrigerant circulation holes, N, to evaluate the performance of the individual refrigerant circulation holes 20.

FIG. 4 shows the simulation results. In the figure, the vertical axis indicates the heat exchange performance index Q/Fa/ΔPa described above, and the horizontal axis indicates the polynomial (W−t1−t2)×Hp×Hf/N.

The figure shows cases where the flat tube width is mm, 14 mm, 15 mm, and 16 mm. As can be seen from the figure, the maximum points of all the curves are contained in the region where the polynomial is 3.95 or greater and 10 or less. Accordingly, setting the polynomial to 3.95 or greater and 10 or less allows a high-performance condenser 1 to be provided.

For comparison with FIG. 4, the polynomial of this embodiment was calculated for the condenser specified in PTL 1. For the specifications of the condenser in PTL 1, the numerical values described in [0021] (tube height th=1.7 mm, fin height Fh=7.8 mm, and outer circumferential tube thickness Td=0.35 mm) and the number of refrigerant circulation holes, 14, shown in FIG. 2 of PTL 1, were referred to.

The diameter of the refrigerant circulation hole was calculated to be 1 mm (=1.7−2×0.35).

Since the flat tube width W is not specified in PTL 1, 16 mm was used as an assumed numerical value to allow comparison with this embodiment. The wall thickness at one end, t1, and the wall thickness at the other end, t2, were individually calculated to be 0.133 mm from the flat tube width, 16 mm, the number of refrigerant circulation holes, 14, and the diameter of the refrigerant circulation hole, 1 mm. The value of the polynomial calculated from the foregoing numerical values is as follows:

$\begin{matrix} {{\left( {W - {t\; 1} - {t\; 2}} \right) \times {Hp} \times {{Hf}/N}} = {\left( {16 - 0.133 - 0.133} \right) \times 1.7 \times {7.8/14}}} \\ {= {14.9.}} \end{matrix}$

This shows that the condenser disclosed in PTL 1 is out of the range of 3.95 or greater and 10 or less, which is the range of the polynomial defined in this embodiment.

FIG. 5 shows the simulation results. In the figure, the vertical axis indicates the heat exchange performance index Q/Fa/ΔPa described above, and the horizontal axis indicates the equivalent diameter de of the plurality of refrigerant circulation holes 20 formed in each of the flat tubes 14. Here, the equivalent diameter de means the diameter when the plurality of refrigerant circulation holes 20 formed in a single flat tube 14 is converted to a single equivalent cylinder.

The figure shows cases where the flat tube width is 12 mm, 14 mm, 15 mm, and 16 mm. As can be seen from the figure, the maximum points of all the curves are contained in the range where the equivalent diameter de is 0.5 or more and 0.8 or less, preferably, 0.55 or more and 0.76. Accordingly, setting the equivalent diameter de as described above allows a high-performance condenser 1 to be provided.

This embodiment provides the following operational advantages.

For evaluation of the heat exchange performance of the condenser 1, the polynomial using the flat tube width W is used. This allows also state changes due to air passing through the corrugated fins 15 (for example, airflow resistance and heat exchange in the airflow direction) to be considered, thus allowing the heat exchange performance to be reflected more rigorously.

The configuration of the heat exchanger is determined in consideration of not only the configuration of the flat tubes, such as the flat tube width W and the flat tube height Hp, but also the fin height Hf. This allows the air pressure loss to be considered more rigorously.

Dividing the polynomial by the number of refrigerant circulation holes, N, allows evaluation of the performance per refrigerant circulation hole.

The performance is evaluated using the exchanged heat Q relative to the frontal surface area Fa and the air pressure loss ΔPa as a heat exchange performance index and using the foregoing polynomial. By evaluating the performance by using the exchanged heat Q divided by the air pressure loss ΔPa, the air pressure loss is sufficiently considered. This allows evaluation of the performance in a state close to the actual operating state.

Since the exchanged heat Q is calculated by also considering the refrigerant pressure loss during the simulation, performance in a state even closer to the actual operating state can be evaluated.

The evaluation using the above polynomial as a parameter showed that the foregoing heat exchange performance index is large (the air pressure loss is small, and the exchanged heat is large) when the value of the polynomial is in the range of 3.95 or greater and 10.0 or less. Thus, a high-performance condenser can be obtained with high reproducibility by defining the high-performance condenser configuration by using the polynomial.

REFERENCE SIGNS LIST

-   1 condenser (heat exchanger) -   13 core portion -   14 flat tube -   15 corrugated fins -   W flat tube width -   t1 wall thickness at one end -   t2 wall thickness at the other end -   Hp height of flat tube in the stacking direction -   Hf height of corrugated fins in the stacking direction -   Pf fin pitch of corrugated fins -   N number of refrigerant circulation holes 

1. A heat exchanger comprising flat tubes, each having therein a plurality of refrigerant circulation holes, and fins fixed to the flat surfaces of the flat tubes, on the surfaces of which air passes, wherein the flat tubes and the fins are alternately stacked to form the heat exchanger, and wherein (W−t1−t2)×Hp×Hf/N is set to 3.95 or greater and 10.0 or less, where de is the equivalent diameter of the plurality of the refrigerant circulation holes, W is the width of each of the flat tubes, t1 is the wall thickness of each of the flat tubes at one end in the widthwise direction corresponding to the distance from the refrigerant circulation hole closest to the one end, t2 is the wall thickness of each of the flat tubes at the other end in the widthwise direction corresponding to the distance from the refrigerant circulation hole closest to the other end, Hp is the height of each of the flat tubes in the stacking direction, Hf is the height of the fins in the stacking direction, and N is the number of the refrigerant circulation holes, and the equivalent diameter de is set to 0.5 or more and 0.8 or less, and the flat tube width W is set to 12 mm or more and 16 mm or less.
 2. The heat exchanger according to claim 1, wherein the equivalent diameter de is set to 0.55 or more and 0.76 or less.
 3. The heat exchanger according to claim 1, wherein the fins have a corrugated shape whose pitch is set to 1.6 mm or more and 2.0 mm or less.
 4. A vehicle air conditioner comprising the heat exchanger according to claim
 1. 